Axial Piston Compressor

ABSTRACT

Compressor, especially a compressor for the air-conditioning system of a motor vehicle, having a housing ( 1 ) and, for drawing in and compressing a coolant, a compressor unit arranged in the housing ( 1 ) and driven by means of a drive shaft, the compressor unit being regulated by means of the pressure (P C ) prevailing in a drive mechanism chamber substantially bounded by the housing ( 1 ), there being an additional regulation device ( 17 ) and/or control device for the inlet-gas-side coolant mass flow and/or the inlet pressure and/or the inlet density.

The invention relates to an axial piston compressor, especially to a compressor for the air-conditioning system of a motor vehicle, in accordance with the preamble of claim 1.

From DE 197 49 727 A1 there is known an example of an axial piston compressor of such a kind, having a housing and, arranged in the housing and driven by way of a drive shaft, a compressor unit for drawing in and compressing a coolant. The compressor unit is regulated substantially by the pressure P_(C) in a drive mechanism chamber depending in the particular instance on the load and/or speed of rotation of the compressor, an inlet pressure P_(V1) and a high pressure P_(V2) that prevail on the inlet and outlet sides, respectively, of the compressor also having an influence on regulation of the compressor. Regulation takes place by way of a change in the piston stroke of the compressor, the piston stroke being governed by the deflection of a tilt plate from a zero position.

In the case of such an axial piston compressor it is additionally the case, however, that because of the force conditions prevailing in the compressor, especially because of centrifugal forces, the compressor has a tendency to up-regulate at high speeds of rotation, that is to say the compressor has a tendency towards a greater piston stroke and accordingly towards a higher pressure on the high-pressure side.

DE 195 14 748 C2 explains the tilting moments which in the case of, in principle, all compressors existing in the prior art of the tilt plate form of construction act on the tilt plate and which decisively contribute to the tilting or, that is to say, inclination behaviour of the tilt plate. The compressor tilting behaviour described in that publication can be regarded as being exemplary of compressors of the tilt plate form of construction. In general it is the moments given below which, in the centre of the tilting movement of the tilt plate, have an influence on the tilting of the tilt plate. The direction of the moment is given in brackets, with (−) denoting down-regulation, that is to say regulation in the direction of minimum stroke, and (+) denoting up-regulation, that is to say regulation in the direction of maximum stroke of the pistons. It is substantially the following moments that play a part:

-   -   moment due to gas forces in the cylinder spaces (+)     -   moment due to gas forces from the drive mechanism chamber (−)     -   moment due to a restoring spring (−) or     -   moment due to an advancing spring (+)     -   moment due to rotating masses (−), these including, for example,         the tilt plate (including a moment due to the position of the         centre of gravity, in which case—for example in accordance with         DE 195 14 748 C2—this component can be positive, that is to say         (+))     -   moment due to masses moved in translation (+), which can         include, for example, pistons, sliding blocks or also an         oscillating wobble plate.

As can be seen from the above listing, the moment due to the rotating masses, hereinafter referred to as M_(SW), generally has a down-regulating action over an extensive tilt angle range. It is only in the region of very small tilt angles, as a result of, for example, an outwardly displaced location of centre of gravity (Steiner component in the calculation of the moment of deviation J_(yz)) that an up-regulating moment can be produced in the case of the tilt plate. DE 195 14 748 C2 also shows a plot of the moment due to masses moved in translation, which as already explained has an up-regulating action.

Also of interest is the sum of the moments, which in the case of the subject-matter of DE 195 14 748 C2 is responsible for up-regulating behaviour of the compressor over the entire tilt angle range of the tilt plate, because the masses moved in translation dominate the regulation behaviour in a wide tilt angle range.

With reference to the prior art, especially according to DE 195 14 748 C2, it is disadvantageous therein that, when there is an increase in the delivery volume due to an increase in the speed of rotation, an additional increase in the delivery volume due to an increase in the tilt angle of the tilt plate is added thereto. This has to be compensated by appropriate interventions for the purpose of regulation, which is onerous, reduces the efficiency of the driving engine and accordingly increases the fuel consumption.

From EP 0 809 027 A1 there is known a compressor in which an attempt is made so to compensate the delivery quantity of the compressor by means of the dynamic behaviour of the compressor drive mechanism that the delivered quantity of coolant or, that is to say, the coolant mass flow can be kept constant. For regulating the delivery quantity so that it is constant in the case of changing speeds of rotation it is proposed that the restoring torque of the wobble plate be utilised, the inclined position of which is counteracted by the restoring torque by virtue of dynamic forces on the co-rotating plate part.

Measures are known from DE 198 39 914 A1 as to how regulation behaviour of such a kind—that is to say, at least partial compensation of the delivery quantity—can be achieved. It is proposed that the tilt plate component mass be so dimensioned in relation to the masses moved in translation that the centrifugal forces of the tilt plate influence the regulation behaviour of the tilt plate. According to DE 198 39 914 A1 it is proposed that the rotating mass of the tilt plate or, that is to say, of the tiltable part of the tilt plate be greater than the total mass of all the pistons so that the centrifugal forces occurring during rotation of the tilt plate are sufficient to counteract the tilting movement of the tilt plate in regulating manner deliberately and thereby to influence, especially to reduce or to limit, the piston stroke and, accordingly, the delivery quantity.

In DE 103 29 393 A1 belonging to the Applicant it is furthermore explained why the component mass should not be the preferred parameter for influencing as desired the regulation behaviour of the drive mechanism resulting from changes in the speed of rotation. According to DE 103 29 393 A1, the desired regulation behaviour of the compressor is primarily achieved not by means of the tilt plate component mass in relation to the masses moved in translation but rather by taking into account the mass moment of inertia of the tilt plate arrangement, which depends more on its geometry than on its component mass. A central concept in that Application is that, in the case of fluctuations or changes in the speed of rotation, the moment due to masses moved in translation is directly compensated, or even over-compensated, by the moment due to rotating masses. In the case of new compressors it is in fact desirable to reduce to a low level the frequency and intensity of interventions for the purpose of regulation.

At this point reference should again be made to EP 0 809 027 A1, in which the objective is set of achieving regulation of the delivery quantity so that it is constant. It can, however, be simply demonstrated that appropriate down-regulation is not possible solely by means of the advancing moment acting on the tilt plate. The delivery volume is directly proportional to the speed of rotation, which is to say, therefore, that when the speed of rotation doubles the delivery volume also doubles. However, for the tilt plate tilt moment, which is produced by the relevant moment of deviation, the following equation applies: M_(SW)=J_(yz)ω². Because it is the square of the speed of rotation that influences the tilt moment, the aim of regulating the delivery quantity so that it is constant cannot be achieved solely by designing or dimensioning an appropriate tilt plate.

The subject-matter of DE 198 39 914 A1 and of DE 103 29 393 A1 already discussed hereinbefore shows approaches for solutions to and arrangements for the objective formulated in EP 0 809 027 A1. In tendential terms it is possible in the case of fluctuations in the speed of rotation, for example therefore in the case of an increase in the speed of rotation at the compressor shaft, for the tilt angle of the tilt plate either to increase (for example, in accordance with DE 195 14 748), which is undesirable, or else in accordance with DE 198 39 914, DE 103 29 393 or EP 0 809 027 A1 to decrease, in which case it should be mentioned that even a decrease or a decreasing tendency in the tilt angle or, that is to say, deflection angle of the tilt plate is of only qualified desirability. In any event, the desired objective is possible only by means of a complicated arrangement which also has to be balanced in highly precise manner especially so that the moment distributions are taken into account. This results in compressors according to the prior art being relatively expensive to manufacture.

In the (unpublished) Patent Application DE 103 47 709 belonging to the Applicant, it is proposed that the active moments due to the mass forces or, that is to say, the moments due to the moments of deviation be so matched that the deflection angle of the tilt plate does not change substantially in the case of changing speeds of rotation. In the case of the subject-matter of DE 103 47 709 it was recognised that regulation behaviour of such a kind represents optimal drive mechanism behaviour in order, by that means, to be able to regulate the mass flow of a coolant compressor in optimal manner.

FIG. 15 shows in diagrammatic form how a compressor 101 of the described form of construction is regulated. Such a compressor makes available, in operation, an inlet gas pressure level and also a high-pressure level. The coolant circuit also has these same pressure levels. A certain pressure modification or, that is to say, pressure adjustment is accomplished by means of an expansion element 103, which in turn reacts to changes in the operating state of the compressor and, where appropriate, intervenes for the purpose of regulation. In the compressor drive mechanism chamber there is established, for example by means of regulatory valves of the compressor, a pressure which is between the inlet gas pressure level and the high-pressure level. The drive mechanism chamber pressure intervenes in the force equilibrium or, that is to say, moment equilibrium at the tilt plate in such a way that the tilt angle of the tilt plate can be modified. If the drive mechanism chamber pressure is set slightly above the inlet pressure, the tilt plate is adjusted to the maximum tilt angle. If the drive mechanism chamber is set substantially above the inlet pressure, the tilt angle is adjusted to the minimum tilt angle. Regulation is accomplished by means of the possible volume flows between the individual chambers or pressure levels. Furthermore, reference numeral 102 denotes a gas cooler/liquefier, reference numeral 104 denotes an evaporator and reference numeral 105 denotes a regulation path. As an alternative to the regulation path shown by the solid line between the gas cooler/liquefier 103 and the pressure level P_(V1), wherein P_(V1) is selected as the desired value, there is shown, by a broken line, a second possibility for the regulation path 105, which has P_(V2) as the desired value. A regulation path of such a kind is more usual especially for the coolant CO₂. The described model is shown here in simplified form and is to be considered as being by way of example.

Because, in operation of the compressor or, that is to say, operation of the vehicle, the speed of rotation of the compressor drive mechanism or, that is, of the vehicle engine changes almost constantly, interventions for the purpose of regulation are constantly required in the case of compressors according to the prior art in order, for example, to ensure a constant delivery output of the compressor or to maintain the desired value. Because the drive mechanism chamber has a comparatively large volume, regulation by means of appropriate modification of the drive mechanism chamber pressure is sluggish and substantial overshoot occurs. Consequently, a constant delivery output of the compressor can be achieved only under difficult circumstances. Likewise, the interventions for the purpose of regulation reduce the performance of the vehicle and consume power of the driving engine.

Starting from the prior art discussed hereinbefore, the objective of the present invention is to provide a compressor (especially, but not exclusively, a tilt plate compressor having variable piston stroke), for use in vehicle air-conditioning systems, which has a speed of regulation that is substantially improved compared to the prior art and which can keep the coolant mass flow constant over wide speed of rotation ranges without considerable loss of performance.

The objective is met in accordance with the invention by a compressor having the features of patent claim 1, advantageous developments and details of the invention being described in the subordinate claims.

It is accordingly a fundamental point of the present invention that a compressor having a compressor unit in accordance with the preamble of patent claim 1 has an additional regulation device and/or control device for the inlet-gas-side coolant mass flow and/or the pressure on the inlet side and/or the inlet density. A constructional measure of such a kind brings about, in addition to throttling of the coolant mass flow, a down-regulating effect (constituting the major part of the total effect), which is produced by utilising the pressure difference—brought about by the throttling—between the pressure side or, that is to say, the pressure (P_(V2)) prevailing on the outlet side and the inlet side or, that is to say, the pressure level P_(V1)* acting on the pistons. For the purposes of illustration, a construction according to the invention applied to an axial piston compressor will be considered hereinbelow, although it should be noted that this must not in any way be interpreted in limiting manner because a construction according to the invention can likewise be applied to a whole series of compressors of some other form of construction. If the coolant flows into the cylinders of the axial piston compressor at a relatively low pressure level or, that is, a relatively low inlet density, the compressor has the tendency, given constant pressure in the drive mechanism chamber, to down-regulate, that is to say to reduce the piston stroke. Accordingly, limitation of the coolant mass flow or, that is, of the inlet pressure intervenes directly in the regulation of the compressor. Expressed in other words, a modification of the volume flow results for the main part in a change in the pressure difference which regulates the compressor and, therefore, in a predetermined regulation characteristic.

In a preferred embodiment of a compressor according to the invention, the regulation device comprises a throttling location having an adjusting member. The throttling location can be, especially, a throttling valve or a throttling flap. Also feasible is a pressure reducer. The adjusting member provides for regulation of the coolant mass flow or, that is, the inlet pressure acting on the piston (P_(V1)*). Such a measure is simple to put into practice in terms of construction and ensures low manufacturing costs.

Optionally, the adjusting member of the regulation device adjusts the coolant mass flow or, that is, the inlet pressure in dependence on the speed of rotation. Speeds of rotation are a readily available regulation variable. Detection of speeds of rotation can be accomplished, for example, by means of the generation of electric pulses (induction principle); however, direct, centrifugal-force-dependent regulation, for example, is also feasible. Accordingly, regulation that is dependent on speed of rotation opens up the possibility of many different forms of construction for a compressor according to the invention, with low manufacturing costs providing an advantage here too.

The throttling location preferably comprises an end-stop, associated with the adjusting member, for a position of minimum flow cross-section, this end-stop being so arranged that, even in the case of very high speeds of rotation of the compressor, a predetermined minimum coolant mass flow or, that is, inlet pressure, is ensured. By this means, it is ensured in simple manner that, even in the case of very high speeds of rotation, the compressor does not automatically down-regulate completely as a result of the throttling location or, that is, the additional regulation device.

In a variant of a compressor according to the invention that is simple and therefore economical to produce, the adjusting member is an adjusting piston, which can especially be in the form a stepped piston.

As an alternative to or in addition to the described regulation device, the said control device is provided for control or, that is to say, limitation of the inlet-gas-side coolant mass flow or, that is, of the inlet pressure or, that is, the inlet density. A control device usually has an arrangement that is simple in terms of construction. In a preferred embodiment, the control device comprises at least one inlet valve arranged on the inlet gas side. The control device can be integrated in the inlet valve(s). Preferably, the inlet valve is a pressure-controlled reed valve, which in a variant that is simple in terms of construction is formed by a valve plate, which has a throttling through-bore, and an inlet blade. The inlet blade is preferably of tongue-like construction.

When the compressor according to the invention is a compressor having pistons, especially an axial piston compressor, which has a cylinder block and at least one, but especially 5 to 9, piston(s), which is/are axially movable back and forth in bores provided in the cylinder block, an inlet valve can optionally be associated with each cylinder, in which case the corresponding inlet blades for the cylinders are integrated in an inlet blade plate. This reduces the number of individual parts required for a compressor according to the invention, which reduces the manufacturing costs. That end of the or each cylinder space which is associated with, or which faces, the inlet valve has, in a further preferred embodiment, an annular extension which can be bevelled off or flattened off towards the fixing location of the inlet blade. By that means the stroke of the inlet blade can be effectively limited.

Again assuming that the compressor is a compressor having pistons, it is advantageous if the ratio of piston diameter and piston stroke is approximately from 0.4 to 1.5, especially approximately from 0.65 to 1.1. The ratio of piston diameter and the throttling through-bore in the valve plate preferably is approximately from 1.5 to 5, especially from 2.5 to 4. The ratio of the throttling through-bore in the valve plate and the stroke of the inlet blade is, in a further preferred embodiment, approximately from 2.5 to 8, especially from 3.7 to 6.7. The ratio of piston stroke to the stroke of the inlet blade can be approximately from 10 to 30, especially from 14 to 24. The above-described ratios, that is to say therefore the geometric characteristics of the above-described compressor, are advantageous in energy terms especially for compressors having CO₂ as coolant. Alternatively or additionally, the control device can also comprise, arranged on the inlet gas side, an orifice plate which appropriately defines the coolant mass flow or, that is, the inlet pressure or, that is, the inlet density.

When the compressor according to the invention is a compressor which has a tilt plate, the deflection angle of the tilt plate, which angle governs the piston stroke of the compressor, is governed to a very large extent by the interaction of, on the one hand, the pressure P_(C) in a drive mechanism chamber substantially accommodating the tilt plate and, on the other hand, the coolant mass flow on the inlet side or, that is, the inlet pressure P_(V1)*. A further force acting on the piston is produced by the pressure P_(V2) on the high-pressure side. As a result of regulation of the pressure in the drive mechanism chamber, on the one hand, and regulation or control of the inlet pressure P_(V1)*, on the other hand, it is accordingly possible for ideal regulation of the piston stroke to be accomplished, with preference being given to modification of the pressure in the drive mechanism compartment or, that is to say, the drive mechanism chamber for “major” interventions for the purpose of regulation, whereas a fine adjustment can be carried out by a rapid regulation or, that is to say, the limitation, defined by the control device, of the inlet pressure. As already mentioned, regulation or control of the inlet pressure is associated with a substantially lower load for the engine than regulation of the pressure in the drive mechanism chamber, so that small rapid interventions for the purpose of regulation can be carried out without major loss of performance or do not even become necessary in the first place.

The regulation device can be actuatable or controllable from outside the compressor. For the purpose there especially comes into consideration a solenoid or like arrangement. This ensures simple maintenance and simple replacement of the actuating device for the regulation device.

In an embodiment which is simple in terms of construction and which therefore is preferred, the regulation device and/or control device include(s) an inlet-gas-side oil separator, which has multifunctional significance. On the one hand, oil present in the inlet gas is separated out; on the other hand, pressure regulation or, that is, regulation of the coolant mass flow can be simultaneously accomplished by that means.

In a variant of a compressor according to the invention, the regulation device is self-regulating and, especially, dependent on the difference in pressures at, on the one hand, the outlet side or, that is to say, the high-pressure side and, on the other hand, the entry side or, that is to say, the inlet side of the compressor. This ensures reliable regulation of the compressor taking into account the most important operating parameters.

Preference is given, especially in the case of a compressor of the tilt plate form of construction, to the moment distribution of those components of the compressor that are moved or movable in rotation and in translation being such that, in the case of an increase in the compressor speed of rotation, a substantially constant regulation characteristic is ensured (that is say, the moments are balanced). Explaining this again using the example of a tilt plate compressor, this means that the tilt angle of the tilt plate remains substantially the same or decreases. Accordingly, in an advantageous combination of features of a compressor according to the invention (again explained using the example of the tilt plate compressor), there are three mechanisms which influence regulation of the compressor, namely firstly the coolant mass flow at the inlet gas side or, that is, the inlet pressure P_(V1)*, which is arranged to be controlled or regulated in accordance with the invention, and also the pressure in the drive mechanism chamber and the moment distribution of the components of the compressor. This means that, by virtue of the form of construction of the tilt plate and/or of the pistons, the moment distribution or, that is, the ratio of moments causes the compressor to behave neutrally in relation to the speed of rotation, that is to say especially it does not up-regulate. Down-regulating action is, if required, assisted by means of an appropriate regulating intervention for the pressure P_(C) in the drive mechanism chamber, with it being possible especially for minor regulating interventions to be accomplished without appreciable loss of engine performance by means of adjustment or, that is to say, defined control/limitation of the inlet pressure P_(V1)*.

The regulation device and/or control device can be arranged in an inlet gas channel extending mainly in the cylinder head of the compressor. The inlet gas channel connects an inlet gas entry of the compressor with an inlet gas chamber which is arranged upstream of or, that is to say, before the inlet openings of the individual cylinders.

The regulation device can furthermore comprise means for measuring the coolant mass flow and/or the pressure in the inlet gas channel (both to each side of the throttling location and also to just one side of the throttling location) and/or the speed of rotation of the compressor and/or of the engine driving the latter and/or the pressure on the outlet side of the compressor. Depending on the specific form of construction of the compressor, regulation thereof is accordingly possible on the basis of a readily available variable.

Optionally, the adjusting member of the regulation device acts against the force of a resilient element, especially against the force of a spring. This makes possible a wide range of different regulation characteristics (depending on which characteristic the resilient element has) and at the same time constitutes automatic regulation of the adjusting member which is simple to put into practice. In a preferred arrangement thereof, the force of the resilient element can be adjusted especially by an adjusting screw or like arrangement. This ensures that, with one and the same construction, various regulation characteristics can be set by simple means. Also, tolerances in the manufacture of the resilient element and/or in the properties of the resilient element can accordingly be compensated by simple means because fine adjustment of the characteristic of the resilient element is possible.

In a form of construction which is especially simple to put into practice in constructional terms, the adjusting member is arranged between the pressure gas side and inlet gas side and is accordingly regulated automatically (where applicable against the action of the resilient element) as a result of being subjected to the pressures of, on the one hand, the pressure gas side and, on the other hand, the inlet gas side.

In a further preferred embodiment, the regulation device and/or control device comprise(s) a throttling location of constant cross-section. This throttling location can be present, for example, as the sole regulating device in a compressor according to the invention or also in combination with a throttling location which has an adjusting member. Especially in the case of compressors with a high working pressure, that is to say, for example, compressors which use CO₂ as coolant, it is possible just by means of this simple constructional measure for the desired outcome to be achieved. Optionally, the inlet line and/or a connection between the compressor and an evaporator are a component part of the regulation device, it being possible, especially in the case of a throttling location of constant cross-section, for efficient regulation of the compressor to be achieved by means of appropriate design of the inlet line and/or of the connection between the compressor and the evaporator.

An especially efficient variant of a compressor according to the invention is produced when across the regulation device there is a pressure difference of approximately 1 bar at a compressor speed of rotation of approximately 600 rpm and/or of approximately 10 bar at approximately 8000 rpm. The regulation device or, that is, the throttling location preferably comprises a tubular line having a tubular cross-section of approximately from 8 mm to 10 mm, which ensures, especially in the case of a throttling location of constant cross-section, a desired regulation characteristic. As already mentioned hereinbefore, CO₂ is used as coolant in a particular form of construction of a compressor according to the invention.

A compressor which is especially efficient and which manages with few interventions for the purpose of regulation relating to the pressure P_(C) in the drive mechanism chamber is produced when the moment due to those components of the compressor that are moved in rotation is of substantially equal magnitude to the moment due to those components of the compressor that are moved in translation, that is to say when the behaviour of the compressor is neutral, in terms of its regulation behaviour, with regard to speed of rotation.

The invention will be described hereinbelow with reference to further advantages and features by way of example and with reference to the accompanying drawings. The drawings show in

FIG. 1 a first preferred embodiment of a compressor according to the invention in a sectional view,

FIG. 2 a diagram of the mode of operation of a second preferred embodiment of a compressor according to the invention,

FIG. 3 a diagram of a co-ordinate system on which the calculation of the moment ratios is based,

FIG. 4 a sectional view and an exploded view of a tilt plate mechanism;

FIGS. 5 to 7 the regulation characteristics of a compressor for various moment ratios of the components of the compressor that are movable in rotation and that are movable in translation,

FIG. 8 a piston of a compressor of the first or second preferred embodiment with the pressure conditions acting on it,

FIGS. 9 a to 9 c a mass flow diagram, a p-V diagram and a regulation characteristic of a compressor according to the invention,

FIGS. 10 to 13 further examples of regulation characteristics of a compressor which has an up-regulating tendency in the case of increasing speed of rotation and of a compressor which exhibits a down-regulating tendency, and

FIG. 14 a detail, in diagrammatic form, of a third preferred embodiment of a compressor according to the invention.

From FIG. 1 it can be seen that a first preferred embodiment of a compressor according to the invention comprises a housing 1, a cylinder block 2 and a cylinder head 3. Pistons 4 are mounted in the cylinder block 2 so as to be movable axially back and forth. The compressor is driven by means of a drive shaft 6, from a belt disc 5. The compressor in this case is a compressor having variable piston stroke, the piston stroke being governed by the deflection angle of a tilt plate 7. The tilt plate 7 is in operative engagement with the pistons 4 by way of sliding blocks 8 and is driven in rotation by the drive shaft 6. The deflection angle of the tilt plate 7 can, as is known from the prior art, be influenced by a pressure change in a drive mechanism chamber 8′, in which the tilt plate is substantially arranged. The drive mechanism chamber 8′ can be subjected to pressures between an inlet pressure, that is to say the pressure prevailing on the inlet side of the compressor, and a high pressure, that is to say the pressure prevailing on the outlet side of the compressor. Depending on the pressure prevailing in the drive mechanism chamber 8′ or, that is, depending on the difference in the pressures on the inlet side and in the drive mechanism chamber, there is produced a predetermined deflection angle for the tilt plate and, accordingly, a predetermined pressure on the outlet side of the compressor.

A second quantity influencing the deflection angle of the tilt plate 7 is the distribution of moments between the components of the cylinder that are movable in translation, for example the piston 4 or the sliding blocks 8, and the components of the compressor that are movable in rotation, for example the tilt plate 7. In this case, by means of an appropriate mass or, that is, moment distribution it is possible to achieve a rather down-regulating tendency for the compressor at high speeds of rotation. This is desirable, particularly in the case of modern compressors, in order to be able to avoid icing-up, especially at high speeds of rotation, without a large number of interventions for the purpose of regulation. More precise details of the exact constructional arrangement relating to the moments will be given following a brief explanation of the further important features of the compressor according to the invention in accordance with the first preferred embodiment.

As can furthermore be seen in FIG. 1, an inlet gas channel 9 is arranged in the cylinder head 3, which channel connects an inlet gas entry 10 to an inlet gas chamber 11, which is arranged upstream of the cylinders. The compressed fluid or, that is to say, coolant is made available to the coolant circuit by way of a pressure gas chamber or, that is to say, outlet chamber 12. For regulation of the coolant mass flow on the inlet gas side and also therefore of the pressure on the inlet side of the compressor, a regulation device is arranged in the inlet gas channel 9. This regulation device comprises an adjusting piston 13 (which as an alternative to the arrangement shown can also be in the form of a stepped piston), a resilient element in the form of a spring 14 and also an adjusting screw 15, which serves to adjust the biasing of the spring 14. On its side facing the outlet chamber or, that is to say, the pressure gas chamber 12, the adjusting piston 13 is subject to the outlet pressure or, that is to say, the high pressure, whereas on its side facing the adjusting screw 15, that is to say on the side facing the inlet gas entry 10, it is subjected to the inlet pressure or, that is to say, the entry pressure. The higher the outlet pressure of the compressor, the further the piston 13, which constitutes the adjusting member of the regulation unit, is pressed into the inlet gas channel 9 and, as a result, narrows the cross-section thereof. This results in a lower coolant mass flow into the inlet gas chamber 11, which leads to a lower pressure in the inlet gas chamber 11 and accordingly in down-regulating behaviour of the compressor.

It is not clearly visible from the drawing that an end-stop for a position of minimum flow cross-section is associated with the adjusting piston 13, thereby ensuring that, even in the case of very high speeds of rotation of the compressor and a relatively high output pressure, a predetermined minimum coolant flow or, that is, inlet pressure is ensured. The regulation unit comprising the adjusting piston 13, the spring 14 and the adjusting screw 15 is accordingly self-regulating, the regulation being accomplished in dependence on the pressures on the outlet side and on the entry side or, that is to say, the inlet gas side. Accordingly it can at this point be stated that the deflection angle of the tilt plate 7 is governed by the interaction of the pressure in the drive mechanism chamber 8′, on the one hand, and the coolant mass flow on the inlet gas side or, that is, the inlet pressure, on the other hand, the inlet pressure itself being dependent in turn on the output pressure of the compressor, so that feedback regeneration is brought about for the compressor.

As an alternative to the automatic regulation shown in FIG. 1, the regulation device which is indicated in the diagram of FIG. 2 in general terms as the throttling location 17 can of course also be regulated by external regulating variables and also by external apparatus, for example a solenoid. FIG. 2 shows that the throttling location 17 or, that is to say, the throttle (adjusting member) is regulated by an external signal 16. This signal can be generated, for example, on the basis of a measurement of the mass flow, of the pressure on the high-pressure side or of a pressure difference between the inlet gas channel and the high-pressure side or, that is to say, of a pressure difference in the inlet gas channel obtained from the different pressures P_(V1) and P_(V1)* across the throttling location 17. Of course, other parameters, for example a speed of rotation or also temperatures or like quantities, are also feasible as the basis for the signal 16.

FIG. 2 also shows a diagrammatic representation of the coolant circuit in an h vs. log p diagram (supercritical process, with CO₂ as coolant) in a representation at the throttling location (δPV).

As already mentioned in the description of FIG. 1, the distribution of moments between the masses of the compressor that are moved in translation, for example the pistons 4, and the masses moved in rotation, which include, for example, the tilt plate 7, also has a regulating effect on the compressor. This moment ratio will be discussed in somewhat greater detail hereinbelow. For illustrative purposes, a simplified derivation to be considered as being given by way of example will be considered for the various moments. In the case of complex geometries, especially of the tilt plate (when the illustrative approach no longer provides satisfactory results), the mass moments of inertia and moments of deviation and also other variables influenced by the geometry and density of the materials can be calculated in simple manner by CAD.

In the simplified, yet illustrative derivation of the mass moments of inertia it is assumed that the centre of gravity of the tilt plate is located at the tilting articulation on the mid-axis of the shaft, that is to say no Steiner component or the like has to be taken into account. For the derivation of the moment of deviation the following mathematical relationships generally apply (the co-ordinate system on which it is based being shown in FIG. 3):

J _(yz)2=−J ₃ cos α₂ cos α₃ −J ₂ cos β₂ cos β₃ −J ₃ cos γ₂ cos γ₃

α₁=0 β₁=90° Direction angles of the x axis γ₁=90° relative to the main inertia axes ξ, η, ζ α₂=90° β₂=ψ Direction angles of the y axis γ₂=90°+ψ relative to the main inertia axes ξ, η, ζ α₃90 ° β₃=90°−ψ Direction angles of the z axis γ₃=ψ relative to the main inertia axes ξ, η, ζ

As mentioned hereinabove, the co-ordinate system used herein can be seen from FIG. 3. The following also applies to a “ring”:

$J_{2} = {J_{\eta} = {\frac{m}{4}\left( {r_{a}^{2} + r_{i}^{2} + \frac{h^{2}}{3}} \right)\mspace{14mu} {and}}}$ $J_{3} = {J_{\zeta} = {\frac{m}{2}\left( {r_{a}^{2} + r_{i}^{2}} \right)}}$

(Note: J₃≈2J₂)

For the moment of deviation, which governs the tilting movement, the following applies:

J _(yz) =−J ₂ cos ψ sin ψ+J ₃ cos ψ sin ψ

Independently of FIG. 3, the following holds true for the moment due to mass forces of the pistons:

$\beta_{i} = {\theta + {2\pi \; \left( {i - 1} \right)\frac{1}{n}}}$ Z_(i) = R ω²tan  α cos  β_(i) F_(m i) = m_(k)z_(i) M(F_(m i)) = m_(k)R cos  β_(i)z_(i) $M_{k,{ges}} = {m_{k}R{\sum\limits_{i = 1}^{n}{z_{i}\cos \; \beta_{i}}}}$

and also the moment M_(SW) due to the moment of deviation:

M_(sw) = J_(yz)ω² $J_{yz} = {\left\{ {{\frac{msw}{2}\left( {r_{a}^{2} + r_{i}^{2}} \right)} - {\frac{msw}{4}\left( {r_{a}^{2} + r_{i}^{2} + \frac{h^{2}}{3}} \right)}} \right\} \cos \; \alpha \; \sin \; \alpha}$ $J_{yx} = {\frac{msw}{4}\sin \; 2\; {\alpha \left( {{3r_{a}^{2}} + {3r_{i}^{2}} - h^{2}} \right)}}$

In the context of the invention, the following moment ratio should be established by structural means for any desired tilt angle or tilt angle range:

M_(SW)≧M_(k,ges) or, preferably, the sub-case M_(SW)=M_(k,ges)

As a result, the following also applies:

$\left\lbrack {{\omega^{z}R^{2}m_{k}\tan \; \alpha {\sum\limits_{i = 1}^{n}{\cos^{2}\beta}}} \leq {\omega^{2}\frac{msw}{24}\sin \; 2\alpha \; \left( {{3r_{a}^{2}} + {3r_{i}^{2}} - h^{2}} \right)}} \right\rbrack$

As already mentioned, the (tilting) moment of the tilt plate due to the associated moment of deviation can be deliberately adjusted by means of various parameters (geometry, density distribution, mass, mass centre of gravity) so that

M_(SW)≧M_(k,ges) or, however, the sub-case M_(SW)=M_(k,ges).

In the context of the equations given, the variables denote the following:

-   θ rotation angle of the shaft (the considerations above and below     being made on the basis of θ=0 for the sake of simplicity) -   η number of pistons -   R distance from piston axis to shaft axis -   ω speed of rotation of shaft -   α tilt angle of tilt ring/tilt plate -   m_(k) mass of a piston including sliding blocks or, that is to say,     pair of sliding blocks -   m_(k,ges) mass of all pistons including sliding blocks -   M_(SW) mass of tilt ring -   r_(a) external radius of tilt ring -   r_(i) internal radius of tilt ring -   h height of tilt ring -   g density of tilt ring -   V volume of tilt ring -   β_(i) angle position of piston i -   z_(i) acceleration of piston i -   F_(mi) mass force of piston i (including sliding blocks) -   M(F_(mi)) moment due to mass force of piston i -   M_(k,ges) moment due to mass force of all pistons -   M_(SW) moment due to advancing moment of tilt ring/tilt plate as a     result of the moment of deviation (J_(yz))

FIG. 4 shows the drive mechanism of the tilt plate form of construction used as the basis, by way of example, for the derivation. In the derivation, the tilt moment M_(SW) due to the moment of deviation J_(yz) of the tilt plate is, in simplified manner, set against the masses moved in translation or, that is, the moment M_(K,ges) produced thereby. Forces and moments of the pins and/or of the gas force support or the like are, in simplifying manner, not included in the calculation scheme, being of subordinate importance.

It can be seen from the mathematical relationships that the influence of the speed of rotation can be reduced out from the equation. Also otherwise included are geometric variables which are in particular relationships to one another and which, including component densities and density distributions, can in principle be so selected that the sum of the moments due to mass forces can be adjusted to zero.

FIGS. 5, 6 and 7 each include a calculation scheme in accordance with the equations used. Also shown, as the calculation result, is the moment equilibrium. In addition, there is shown a (qualitative) tilt characteristic, as would result if the gas forces were taken into account.

The tilt characteristics of FIGS. 5, 6 and 7 result when, in addition to the variation of speed of rotation and drive mechanism chamber pressure in addition to the described forces and moments, a particular inlet pressure and a particular high pressure are imposed for system-related reasons. In the process it is assumed that the inlet pressure prevailing upstream of the compressor and the high pressure prevailing downstream of the compressor approximately correspond to the inlet pressure and the high pressure in the compressor, that is to say no throttling takes place in the compressor. With regard to the moment equilibriums calculated in accordance with the given equations there is obtained according to

FIG. 5 a drive mechanism having up-regulating behaviour

FIG. 6 a drive mechanism having down-regulating behaviour, and

FIG. 7 a drive mechanism having neutral behaviour.

Using FIG. 7 and also the equation for the sum of moments, the influence of the tilt angle can be analysed in simple manner. The effect results from the plots of the terms tan(α) and sin(2α). This means that in the calculation the sum of moments can be balanced for precisely one tilt angle; if this is done, for example, for the maximum tilt angle of the tilt plate, there are relatively small deviations in the sum of moments for other tilt angles. It is, however, possible to keep these deviations relatively small.

It is feasible to set the moment equilibrium for the following tilt angles:

for α_(min)<=α<=α_(max): M_(K,ges)=M_(SW) for α=(α_(min)−α_(max))/2: M_(K,ges)=M_(SW) for α_(max)=α_(max): M_(K,ges)=M_(SW) for α>=α_(max): M_(K,ges)=M_(SW)

The two last-mentioned cases are to be preferred.

The advantage of a drive mechanism which is substantially balanced in terms of its sum of moments lies in the fact, inter alia, that when the speed of rotation increases the piston stroke does not increase as well, which is to say that in such a (undesirable) case there would be two effects present which would have to be compensated. It can accordingly be stated that the case is to be preferred where M_(k,ges) is approximately equal to M_(SW), which results in the compressor's having regulation behaviour which is neutral with respect to speed of rotation. If desired, M_(SW) can also be selected so as to be greater than M_(k,ges), which results in down-regulating behaviour of the compressor at high speeds of rotation; on no account, though, is the case desirable where M_(k,ges) is greater than M_(SW) (up-regulation of the compressor in the case of increasing speed of rotation).

As mentioned hereinbefore, it is preferred for M_(k,ges) to correspond approximately to M_(SW). As can be seen from FIG. 7, in a plot of the drive mechanism chamber pressure against the tilt angle, the course of the curves is very similar for all speeds of rotation n, with approximate equality of moments. This is also reflected in a plot of the moment against the tilt angle, from which it can be seen that for all tilt angles the sum of the moments is almost constant. The individual moments do certainly vary, however, for different tilt angles, with M_(k,ges) increasing for greater tilt angles in the entire range shown whereas M_(SW) decreases for greater tilt angles, resulting in the sum of moments M_(k,ges)+M_(SW) shown, which is approximately constant. Accordingly, a compressor characterised by a moment plot of such a kind is, in terms of its regulation characteristic, almost independent of the speed of rotation.

If the effect due to the sum of moments has a down-regulating action, at least the tendency is the right one. However, the influence of the speed of rotation on the effective moments M_(SW) and M_(K,ges) is quadratic, compared to the linear influence of the speed of rotation on the stroke volume and is accordingly only of qualified suitability for keeping the delivered mass flow constant.

Given a drive mechanism that behaves neutrally in the case of changes in the speed of rotation, it is substantially only as a result of change in the pressures P_(V1) (inlet pressure), P_(V2) (high pressure or, that is to say, outlet pressure) and the drive mechanism chamber pressure P_(C) that the tilt angle of the tilt plate changes. At a constant operating point where P_(V1) and P_(V2) are as prespecified, change occurs substantially only as a result of the drive mechanism chamber pressure P_(C).

When a drive mechanism is designed in accordance with the described criteria, the behaviour in relation to the delivered coolant mass flow is proportional when there is a change in the speed of rotation. This means that if the compressor speed of rotation doubles, with the tilt plate tilt angle remaining the same, which is the case with a drive mechanism having neutral behaviour, then approximately double the amount of coolant is delivered. Delivery of precisely double the amount of coolant is the result if further losses that occur as a result of changed flow conditions etc. are disregarded. If the changed flow conditions are taken into account, discrepancies may occur.

In order to keep the delivered coolant mass flow constant, or to limit it, in the case of a, for example, substantial increase in the speed of rotation, there is provided, as already described in the description of FIGS. 1 and 2, in the region of the coolant inlet, a throttling location which is variable and which provides rapid intervention.

It is possible for the compressor to be so designed that the throttling intervenes in direct dependence on the compressor speed of rotation (as, for example, in the second preferred embodiment; see FIG. 2). In the first preferred embodiment (see FIG. 1), on the other hand, the cross-section of the throttling location is a function of the high pressure P_(V2) of the compressor, that is to say the throttling is controlled in dependence on the high pressure.

When the compressor speed of rotation increases (for example, suddenly), then the pressure P_(V2) increases approximately just as quickly. Because no substantial throttling occurs on the high-pressure side, P_(V2) can be assumed to be both the system-side high pressure and also the pressure level on the high-pressure side in the cylinder head. Pressure losses in the pipework play just a subordinate part so that they can be disregarded in this analysis. In the case of the mentioned increase in the compressor speed of rotation the inlet pressure moreover decreases, with the pressure P_(V1), which prevails upstream of the throttling location, approximately maintaining its level (system-side pressure level on the inlet side), whereas the pressure P_(V1)* downstream of the throttling location drops compared to P_(V1). The pressure P_(V2) now acts, as a fundamental adjustment variable (in addition to the inlet pressure), on the throttling mechanism in such a way that the cross-section of the throttling location is reduced.

Lowering of the pressure P_(V1) to the pressure P_(V1)* downstream of the throttling location has the consequence that a reduced inlet density (reduced pressure) is applied to the cylinders (at the inlet valves); as a result, the pressure in the cylinder or, that is, at the end faces of the pistons (which are directed towards the valve plate) decreases so that the tilt angle has a tendency to reduce. This moreover results in the fact that the pressure P_(V2) reduces again, which in turn results in feedback to the throttle.

Because the inlet condition of the compressor, which can be described by the variables t_(V1) and P_(V1) (see FIG. 14), is substantially unchanged, the mentioned expansion valve will not change its setting and the pressure levels P_(V1) and P_(V2) also remain the same. It should be mentioned at this point that the regulation path can also be arranged differently so that, instead of P_(V1), the pressure P_(V2) can be used for defining the state of the compressor.

To summarise, the thermodynamic variables before the compressor and after the compressor (in the direction of circulation) remain the same and the regulation element does not intervene in the system.

In addition to a thermostatic expansion element, differently operating and differently actuated expansion elements are of course also feasible.

As a result of the increase in the compressor speed of rotation, the compressor regulates itself automatically by means of the fact that, in addition to the drive mechanism chamber pressure P_(C), the inlet pressure P_(V1) or the high pressure P_(V2) come to have a regulating effect. Because P_(V1) and P_(V2) are also affected by the operating state of the system, where usually it is not necessarily desirable also to have changes in the case of a changed compressor speed of rotation, a pressure P_(V1)* is brought about which, being the gas force applied to the pistons, can intervene in the force equilibrium or, that is, moment equilibrium of the tilt plate.

This means that, in an operating state with a suddenly increasing speed of rotation, the inlet pressure after the throttling location is reduced so that the pressure level P_(V1)* is established, in order to keep the high pressure P_(V2) and the mass flow at the same level. Acting on the pistons are, on one side, the pressures P_(V1)* and P_(V2) and also, on the other side (on the drive mechanism side), the pressure P_(C) (see FIG. 8). When P_(V1) is lowered to the level P_(V1)*, the tilt angle of the tilt plate is reduced, that being the case without having to change the crank chamber pressure. This means that, in contrast to the prior art, where the drive mechanism chamber pressure P_(C) is used as an adjusting variable, a further adjusting variable P_(V1)* is introduced in accordance with the invention.

The pressure P_(V1)* can be substantially smaller than P_(V1) (certainly by 5 to 15 bar). Because such throttling can, depending on the operating point, be associated with substantial losses, the throttling location or, that is to say, the regulation device is variable over a wide range.

In a preferred embodiment, the throttling location, which depending on its position constricts the inlet gas line cross-section to a greater or lesser extent, has three different operating ranges:

-   -   In the first position, no throttling is brought about (operating         position 1).     -   In the second position, acting on the inlet side on the piston         are a pressure between the pressures P_(V1) and P_(V1)* and on         the high-pressure side the pressure P_(V2). Also acting as a         guiding variable is, for example, the force of a pressure         spring. In the second position, substantial or less substantial,         depending on the gas forces applied, throttling occurs         (operating position 2).     -   In a third position, the adjusting piston can, when the flow         cross-section in the inlet line has reached a minimum, hit an         end-stop. In that event, a minimum flow cross-section is         maintained (operating position 3).

It is to be understood that the principle proposed here is to be regarded as being only by way of example. There can also be used, for example, a stepped piston, where for the pressures P_(V1) and P_(V2) there is available in each case a different application diameter. It should be mentioned at this point that the adjusting member or, that is to say, the piston should operate in leakage-free manner as far as possible, which is ensured by means of piston rings. Other sealing measures are also feasible.

FIG. 9 a shows, for a prespecified pressure level P_(V1) and P_(V2) of the air-conditioning system, the adjustable mass flows (qualitative representation), whilst FIG. 9 b shows the p-V diagram corresponding thereto. Starting from the origin, the achievable mass flow increases along with the speed of rotation. The envelope curve for the corresponding slope shows the delivered mass flow for a maximum tilt plate tilt angle/maximum geometric stroke volume. The coolant mass flow m₁ at the speed of rotation n₁ doubles, for example, to a coolant mass flow m₂ in the event of a corresponding increase in the speed of rotation of n₂=2×n₁. The greater the desired coolant mass flow, the greater the compressor speed of rotation also has to be in order to achieve that flow at maximum compressor stroke. Once the desired coolant mass flow has been achieved, for example m₁, m₂ or m₃, no increase in the coolant mass flow is desired in the event of further increase in the speed of rotation. The horizontal plots for m₁, m₂ and m₃ shown in the diagram correspond in each case to a particular drive mechanism chamber pressure, which is approximately constant. In the region of the horizontal lines, the effect of the inlet-gas-side throttling location comes into play with increasing speed of rotation. Whereas on the envelope curve the inlet-gas-side throttle is in operating position 1 (no throttling), in operating range 2 the throttling cross-section is reduced with increasing speed of rotation.

Consequently, when the throttling location is appropriately designed for various coolant mass flows, which are established by virtue of a particular operating state, the mass flow can be kept constant.

When, for example, the operating state is established, by means of P_(V1), t_(V1) and P_(V2), for the pressures on the high-pressure side and on the inlet side of the system and also for the inlet state at the compressor entry and, in that state of operation, the speed of rotation n₂ is present with a coolant mass flow m₁, the throttling location is in the operating state 2, that is to say the inlet cross-section of the inlet line is reduced in the region of the throttling location with respect to the initial state (operating state 1). Furthermore, in addition to the pressure level at the compressor entry P_(V1), an inlet pressure P_(V1)* will have been established, which because of the throttling is lower than the pressure P_(V1). On further reduction of the pressure P_(V1)*, the gas forces acting on the piston become lower so that, with the drive mechanism chamber pressure remaining approximately the same, the tilt angle of the tilt plate is reduced (in contrast to the prior art; see FIG. 9 c in this regard). A reduction in the tilt plate tilt angle leads in turn to a lower mass flow. In this case, it is substantially not by means of the fact that a pressure loss is produced or that the volumetric efficiency or, that is to say, degree of filling becomes poorer that the delivered amount is limited or kept constant; mainly the pressure reduction intervenes directly in the equilibrium of stroke adjustment and down-regulates the stroke in the case of increasing speed of rotation. It should be mentioned at this point that P_(V1)* should not drop too substantially, because otherwise excessive losses are brought about.

The regulation behaviour is especially characterised in that, in contrast to the prior art, where given a constant operating state of the system described by P_(V1), t_(V1) and P_(V2) exactly one tilt plate tilt angle is associated with each drive mechanism chamber pressure P_(C) (see FIG. 5, although exceptions occur in the case of very high speeds of rotation or very small tilt angles (maxima), a plurality of tilt plate tilt angles are feasible for a drive mechanism chamber pressure P_(C). In contrast to the prior art, not only is P_(C) an adjusting quantity but so is the pressure P_(V1)* too.

The pressure difference P_(V1)*—P_(C) can reach negative values. In the case of the prior art P_(C)-P_(V1) must be used as the basis. The pressure P_(C) is in that case always higher than the pressure P_(V1). As a result, the regulation range is also greater (Δp) in accordance with the invention.

In conclusion it must be mentioned again that in addition to the adjusting variable acting on the adjusting piston (the coolant mass flow m or, that is, P_(V2)), external actuation of an adjusting piston or throttling device can also be carried out (by means of a solenoid or the like; see FIG. 2). Such an arrangement must be “informed” of the mass flow increase in the form of a signal, for example by detecting the inlet-side or high-pressure-side pressure difference (throttling location/measurement orifice (variable or non-variable) on the inlet side or high-pressure side of the compressor).

FIGS. 10 to 13 show continuations of the qualitative representations of FIG. 6 and FIG. 7, with a drive mechanism that is independent of the speed of rotation being shown in FIGS. 10 and 11 and with a drive mechanism that, analogously to FIG. 7, favours down-regulation in the case of increasing speed of rotation being shown in FIGS. 12 and 13. It is shown that in the case of an unchanged sum of moments (ratio of M_(SW) and M_(K,ges) from FIGS. 6 and 7) throttling on the inlet side (formation of P_(V1)* as opposed to P_(V1)) which is dependent on mass flow, pressure or speed of rotation brings about further separation of the characteristic curves.

Finally, FIG. 14 shows, in diagrammatic form, a third preferred embodiment of a compressor according to the invention. The third preferred embodiment is a compressor which does not have a regulation device but rather a control device for the inlet pressure. This results in the fact that the described compressor is very simple in terms of construction and also, therefore, economical to manufacture. It should be pointed out at this point that the control device of the third preferred embodiment can be put into practice in a compressor together with a regulation device for the inlet pressure. Alternatively, however, a construction which has only a control device is also feasible. The third preferred embodiment comprises, as do the other preferred embodiments as well, a plurality of pistons 4, which are mounted in the cylinder block 2 so as to be movable back and forth.

The third preferred embodiment has, instead of a regulatable throttling device, a valve plate 18 having an inlet blade 21 arranged below it, at the inlet side for the inlet gas into the cylinder space. The inlet blade 21 is of tongue-like construction and serves for control of the inlet gas entry. When the gas is compressed in the cylinder, the inlet blade 21 closes a throttling through-bore 19, whereas when the inlet gas is being drawn in (brought about by the lower pressure prevailing in the cylinder) the inlet blade 21 moves in a downward direction through a stroke t (indicated by arrows 20) and allows the coolant or, that is, the inlet gas being drawn in to enter the cylinder through the throttling through-bore 19.

The throttling through-bore 19 has a diameter d. By virtue of the geometry of the inlet valve, that is to say especially by virtue of the diameter d of the throttling through-bore 19, and the compressor geometry, desirable lowering of the pressure P_(V1)* is brought about over wide operating ranges of the compressor according to the invention. This can be achieved, for example, (in the case of a compressor having CO₂ as coolant) using the following parameters for the compressor geometry: The stroke t of the inlet blade 21 is between 0.9 and 1.2 mm, whereas the valve plate 18 has a bore (throttling through-bore) whose diameter d is between 4.5 and 6 mm. The values for the piston diameter are approximately from 15 to 18 mm and the piston stroke is approximately from 17 to 22 mm. The maximum stroke volume per cylinder is from 3 ccm to 6 ccm. This results in variables describing the geometry of the compressor that are advantageous in energy terms which are a ratio of piston diameter and piston stroke of approximately from 0.65 to 1.1, a ratio of piston diameter and the throttling through-bore in the valve plate of approximately from 2.5 to 4, a ratio of the throttling through-bore in the valve plate and stroke of the inlet blade of approximately from 3.7 to 6.7 and a ratio of piston stroke to the stroke of the inlet blade of approximately from 14 to 24. It should be noted at this point that these values reflect the optimum geometry for operation with CO₂ as coolant but that, depending on constructional requirements, values of from 0.4 to 1.5 for the ratio of piston diameter and piston stroke and values of from 1.5 to 5 for the ratio of piston diameter and the throttling through-bore and values of from 2.5 to 8 for the ratio of the throttling through-bore and stroke of the inlet blade and values of approximately from 10 to 30 for the ratio of piston stroke to the stroke of the inlet blade are also advantageous in energy terms. Accordingly, in this preferred embodiment, the throttling through-bore 19 on the inlet side is used as a throttling location or, that is to say, control device and is appropriately designed in conjunction with the other parameters regulating the compressor. A construction of such a kind is, especially, very effective in compressors which are already optimised in terms of moments, that is to say which have an optimum relationship between the moments due to the rotary masses and due to the masses moved in translation. The gas flowing in flows through an inlet chamber, which is located in the cylinder head 2, at a pressure P_(V1) and is then, by way of the inlet valve, which has, for example, the configuration described above, introduced into the cylinder bore, where by virtue of the inlet valve configuration the pressure P_(V1)* is established, which ensures optimum regulation behaviour of the compressor.

In summary it should be stated that the throttling of the inlet pressure or, that is, of the coolant flow produces a down-regulating effect which primarily results not from a lowering of the inlet density but rather from the direct utilisation of the prevailing pressure difference at the throttle for the purpose of stroke volume adjustment. Adjustment of the throttle results in adjustment of the pressure difference prevailing at the throttle and, therefore, in adjustment of the stroke volume. Furthermore, a modification of the volume flow results in a change in the prevailing pressure difference and, therefore, in subsequent regulation of the stroke volume.

It can furthermore be stated that the advantages of the invention lie, inter alia, in the fact that the tilt plate compressor

-   -   reacts less sensitively, or hardly reacts, to variations in the         speed of rotation caused by the belt drive (drive mechanism)     -   the losses due to interventions for the purpose of regulation         between the pressure levels inlet pressure, high pressure and         drive mechanism chamber pressure are reduced     -   the speed of regulation is improved     -   the coolant mass flow can be kept constant in a wide speed of         rotation range, and     -   the coolant mass flow can be limited at high speeds of rotation.

Although the invention is described using embodiments having fixed combinations of features, it nevertheless also encompasses any further feasible advantageous combinations of those features, as are especially but not exhaustively mentioned in the subordinate claims. All features disclosed in the application documents are claimed as being important to the invention insofar as they are novel on their own or in combination compared with the prior art.

LIST OF REFERENCE NUMERALS

-   1 housing -   2 cylinder block -   3 cylinder head -   4 piston -   5 belt disc -   6 drive shaft -   7 tilt plate -   8 sliding block -   8′ drive mechanism chamber -   9 inlet gas channel -   10 inlet gas entry -   11 inlet gas chamber -   12 output gas chamber or pressure gas chamber -   13 adjusting piston -   14 spring -   15 adjusting screw -   16 external signal -   17 regulation device -   18 valve plate -   19 bore -   20 arrow -   21 inlet blade -   101 compressor -   102 gas cooler/liquefier -   103 expansion element -   104 evaporator -   105 regulation path 

1. A compressor comprising a housing (1) and, for drawing in and compressing a coolant, a compressor unit arranged in the housing (1) and driven by a drive shaft, the compressor unit being regulated by a pressure (P_(C)) prevailing in a drive mechanism chamber substantially bounded by the housing (1), and an additional regulation device (17) and/or control device for an inlet-gas-side coolant mass flow and/or an inlet pressure and/or an inlet density.
 2. Compressor according to claim 1, wherein the regulation device (17) comprises a throttling location having an adjusting member (13) comprising a throttling valve or a throttling flap, or comprises a pressure reducer.
 3. Compressor according to claim 2, wherein the adjusting member (13) of the regulation device adjusts the coolant mass flow or, that is, the inlet pressure in dependence on a speed of rotation.
 4. Compressor according to claim 2, wherein the throttling location has an end-stop, associated with the adjusting member (13), for a position of minimum flow cross-section such that even in the case of very high speeds of rotation a predetermined minimum coolant mass flow or inlet pressure is ensured.
 5. Compressor according to claim 2, wherein the adjusting member is an adjusting piston (13).
 6. Compressor, especially according to claim 1, wherein the control device comprises at least one inlet valve arranged on an inlet gas side.
 7. Compressor according to claim 6, wherein the inlet valve is a pressure-controlled reed valve.
 8. Compressor according to claim 6, wherein the inlet valve comprises a valve plate (18) having a throttling through-bore (19) and a tongue-like inlet blade (21).
 9. Compressor according to claim 8, further comprising a cylinder block and at least one piston which is arranged to move axially back and forth in a corresponding at least one bore provided in the cylinder block, the inlet valve comprises at least one inlet valve so that a separate one of the inlet valves is associated with each cylinder and corresponding inlet blades (21) are integrated in an inlet blade plate.
 10. Compressor according to claim 9, wherein an end of each cylinder space which is associated with a respective one of the at least one inlet valve has a radially extending annular extension, which limits a stroke of the associated inlet blade (21) and which is bevelled off or flattened off towards a fixing location of the associated inlet blade.
 11. Compressor according to claim 6, further comprising a cylinder block and at least one piston, which is arranged to move axially back and forth in respective bores provided in the cylinder block, a ratio of piston diameter and piston stroke (D/s) is approximately from 0.4 to 1.5.
 12. Compressor according to claim 11, wherein a ratio of piston diameter and a throttling through-bore in a valve plate (D/d) is approximately from 1.5 to
 5. 13. Compressor according to claim 12, wherein a ratio of the throttling through-bore in the valve plate and a stroke of an inlet blade (d/t) is approximately from 2.5 to
 8. 14. Compressor according to claim 13, wherein a ratio of piston stroke to the stroke of the inlet blade (s/t) is approximately from 10 to
 30. 15. Compressor according to claim 6, wherein the control device is defined by the geometry of the inlet valve.
 16. Compressor according to claim 1, wherein the control device comprises an orifice plate arranged on a inlet gas side.
 17. Compressor according to claim 1, wherein the compressor unit comprising pistons (4) moving axially back and forth in a cylinder block (2) and, driving the pistons (4) and rotating together with the drive shaft (6), a tilt plate (7), especially a swash plate or wobble plate or a tilt ring, a deflection angle of the tilt plate (7) is governed by interaction of, on the one hand, a pressure in a drive mechanism chamber (8′) substantially accommodating the tilt plate and, on the other hand, the coolant mass flow on the inlet side or the inlet pressure.
 18. Compressor according to claim 1, wherein the regulation device (17) is arranged to be actuated or controlled from outside the compressor.
 19. Compressor according to claim 1, wherein the regulation device (17) and/or control device include(s) an inlet-gas-side oil separator.
 20. Compressor according to claim 1, wherein the regulation device (17) is self-regulating in dependence on a difference in pressures at, on the one hand, an outlet side or high-pressure side and, on the other hand, an entry side or inlet-pressure side.
 21. Compressor according to claim 17, wherein a moment distribution of components of the compressor that are moved in rotation and in translation is such that, in the case of an increase in a compressor speed of rotation, the deflection angle of the tilt plate (7) remains substantially the same or decreases.
 22. Compressor according to claim 17, wherein the regulation device and/or control device is/are arranged in an inlet gas channel (9) extending mainly in a cylinder block (2).
 23. Compressor according to claim 22, wherein the regulation device (17) comprises a coolant mass flow measuring device and/or a pressure sensor in the inlet gas channel (9) to one side or to each side of the throttling location and/or the speed of rotation of the compressor and/or of an engine driving the latter and/or a pressure on an outlet side of the compressor.
 24. Compressor according claim 23, wherein the adjusting member (13) of the regulation device (17) acts against a force of a resilient element.
 25. Compressor according to claim 24, wherein the force of the resilient element or biasing exerted by the resilient element on the adjusting member (13) is adjustable by an adjusting screw (15).
 26. Compressor according to claim 25, wherein the adjusting member (13) is arranged between the pressure-gas side and the inlet gas side.
 27. Compressor according to claim 17, wherein the regulation device comprises a throttling location of constant cross-section.
 28. Compressor according to claim 27, wherein an inlet line and/or a connection between the compressor and an evaporator are part of the regulation device.
 29. Compressor according to claim 28, wherein across the regulation device there is a pressure difference of approximately 1 bar at a compressor speed of rotation of approximately 600 rpm and/or of approximately 10 bar at approximately 8000 rpm.
 30. Compressor according to claim 28, wherein the regulation device or the throttling location comprises a tubular line having a tubular cross-section of approximately from 8 to 10 mm.
 31. Compressor according to claim 30, wherein CO₂ is used as coolant.
 32. Compressor according to claim 17, wherein a moment due to components of the compressor that are moved in rotation M_(SW) is of substantially equal magnitude to a moment M_(k,ges) due to components of the compressor that are moved in translation. 